1. Field of the Invention
The present invention relates to a turbine device for use in a power generation plant or the like.
2. Description of the Related Art
Gas turbines and steam turbines have been used to convert the thermal energy of high-temperature gases and steam into mechanical power or electric power. In recent years, it is very important for turbine manufacturers to increase the performance of turbines as energy transducers for preventing energies from being exhausted and also preventing the global warming on the earth.
High- and medium-pressure turbines have a relatively small ratio of the blade height to the inner diameter of the turbine. Therefore, these turbines suffer a large loss due to a secondary flow because of a large effect of a region referred to as a boundary layer where the energy of a fluid developed on inner- and outer-diameter surfaces of the turbine is small. The mechanism of generation of the secondary flow is as follows:
As shown in FIG. 1 of the accompanying drawings, a flow G flowing into a space between two adjacent rotor blades 1 is subjected to a force caused by a pressure gradient from a pressure surface F of one of the rotor blades 1 toward a suction surface B of the other rotor blade 1. In a main flow spaced from an inner-diameter surface L and an outer-diameter surface M (hereinafter referred as to hub endwall and tip endwall), the force caused by the pressure gradient and a centrifugal force caused by the deflection of the flow are in balance. However, flows within boundary layers near endwalls are of small kinetic energy and hence are carried from the pressure surface F toward the suction surface B under the force due to the pressure gradient as indicated by the arrows J. In the latter part of their path, these flows collide with the suction surface B and turn up to form two vortices W. The vortices W cause a low-energy fluid to be accumulated in the boundary layers near the endwalls, producing an non-uniform flow distribution having two loss peaks downstream of the blades, as shown in FIG. 2 of the accompanying drawings. While the non-uniform flow is finally mixed out to uniform downstream of the blades, it brings about a large energy loss.
It has been proposed to suppress the above secondary flow for increasing turbine performance by providing an inclined or curved surface across the entire blade height. However, controlling the secondary flow according to the proposal is not effective unless the blades are largely inclined or curved, and the largely inclined or curved blades often result in a problem in terms of mechanical strength especially if the blades are rotor blades.
Heretofore, high- and medium-pressure turbines have been designed two-dimensionally. With the development of computers and flow analysis technology, however, three-dimensional blade configurations are made applicable to those high- and medium-pressure turbines. The three-dimensional blade configurations make it possible to perform three-dimensional control on a loading distribution on blades which is given as the pressure difference between the pressure and suction surfaces of blades, and to reduce an energy loss of the blades. According to the conventional three-dimensional blade design, a plurality of twodimensional blade profiles at a certain blade height are designed and stacked along the blade height, thus defining three-dimensional blades. Consequently, it is not possible to control the pressure distribution in detail on the blades fully across the blade height for reducing an energy loss.
It is therefore an object of the present invention to provide a turbine device having blades whose loading distribution is three-dimensionally controlled for reducing an energy loss.
According to the present invention, there is provided a turbine device comprising a rotor having a plurality of turbine blades disposed between an inner-dimeter surface and an outer-diameter surface, the turbine blades being of a front or intermediate loaded type near the inner-diameter surface and of a rear loaded type near the outer-diameter surface.
Specifically, the turbine blades are of the front or intermediate loaded type near the inner-diameter surface and of the rear loaded type near the outer-diameter surface by three-dimensionally imparting a distribution of rates of change of circumferential velocity in the turbine blades.
Details of how the present invention has been made will be described below.
The inventors have focused on how best results can be achieved by finding such a position in the meridional direction in a flow path defined by turbine rotor blades, that the turbine rotor blades receive the greatest energy from the fluid, i.e., a position for the greatest load on the turbine rotor blades, at different blade heights. For an easier analysis, the flow path is divided into a front zone, an intermediate zone, and a rear zone along the meridional direction.
Work done by the turbine rotor blades is given as a change in a circumferential component Vxcex8 of the absolute velocity at the rotor blade inlet and outlet, as shown in FIG. 3 of the accompanying drawings. The change in the circumferential component Vxcex8 between the rotor blades is related to a loading distribution that is given as a pressure difference or enthalpy difference between pressure and suction surfaces of the rotor blades, according to the following equations:
For an incompressible flow:
Loading distribution=Ppxe2x88x92Ps=(2xcfx80/B)xcfx81W(∂rxc2x7Vxcex8/∂m)
For a compressible flow:
Loading distribution=hpxe2x88x92hs=(2xcfx80/B)W(∂rxc2x7Vxcex8/∂m)
where Pp, Ps represent static pressure respectively on the pressure and suction surfaces, hp, hs static enthalpy respectively on the pressure and suction surfaces, B the number of rotor blades of the turbine device, xcfx81 the fluid density, W the average value of speeds on the pressure and suction surfaces, and (∂rxc2x7Vxcex8/∂m) the rate of change of the circumferential velocity Vxcex8 between the rotor blades with respect to the axial distance m. These equations indicate that the loading distribution on the turbine rotor blades is related to the rate of change of the circumferential velocity, and that the loading distribution can be controlled by the value of the rate of change of the circumferential velocity. Specifically, if the rate of change of the circumferential velocity increases at an arbitrary position between the rotor blades, the blade surface load (Ppxe2x88x92Ps) or (hpxe2x88x92hs) increases at that position.
Therefore, the blade loading is related to the rate of change of the circumferential velocity in the axial direction of the turbine rotor blades according to the above equations. If the positive direction of the circumferential component Vxcex8 is defined as the direction in which the rotor blades rotate, then since the circumferential component Vxcex8 decreases from the rotor blade inlet toward the rotor blade outlet in the flow path between the rotor blades, the rate of change of the circumferential component Vxcex8 becomes a negative value. FIG. 4 of the accompanying drawings shows a distribution of rates of change of the circumferential component between the turbine rotor blades. Since, in general, the rate of change of the circumferential component decreases in a certain range from the rotor blade inlet, is substantially constant in an intermediate range, and increases in a rear range, there are two boundary values A, B (hereinafter referred to as branch control points) on the distribution. As shown in FIG. 5 of the accompanying drawings, a distribution of rates of change of the circumferential component where two branch control points A1, B1 are present in a front zone of the flow path in the meridional direction is referred to as a front loaded type, a distribution of rates of change of the circumferential component where a first branch control point A2 is present in the front zone of the flow path in the meridional direction and a second branch control point B2 is present in a rear zone of the flow path in the meridional direction is referred to as an intermediate loaded type, and a distribution of rates of change of the circumferential component where two branch control points A3, B3 are present in the rear zone of the flow path in the meridional direction is referred to as a rear loaded type.
When certain loading distributions (front, intermediate, and rear loaded types) were fixed in a mid-span and a tip of rotor blades, effects of loading distributions at a base of rotor blades as they were set to the front, intermediate, and rear loaded types as shown in FIG. 5 were inspected. Blades which are designed based on these loading distributions have cross-sectional profiles at their bases, as shown in FIG. 6 of the accompanying drawings. A computerized flow analysis of flows between turbine rotor blades whose bases have such cross-sectional profiles indicates that velocity vectors near the bases of the turbine rotor blades, i.e., at the inner-diameter surfaces thereof, are as shown in FIG. 7 of the accompanying drawings. It can be seen from FIG. 7 that a flow separation occurs in the middle of the flow path between the blades of the rear loaded type. The flow separation produces a strong secondary flow from the pressure surface toward the suction surface. As shown in FIG. 8 of the accompanying drawings, an energy loss peak near the inner-diameter surfaces (or hub endwall surfaces) of the blades of the rear loaded type is greater than that of the front or intermediate loaded type. No significant difference exists between the energy loss peaks on the inner-diameter surfaces of the blades of the front and intermediate loaded types.
As shown in FIG. 9 of the accompanying drawings, if loading distributions are set to the front loaded type and the rear loaded type at the tip of the blades and the blades are designed based on such loading distributions in the same manner as described above, then the blades have cross-sectional profiles at their tip as shown in FIG. 10 of the accompanying drawings. When certain loading distributions (front, intermediate, and rear loaded types) were fixed in a mid-span and a base of rotor blades, loss distributions at the blade outlet of the blades of the front and rear loaded types at their tip were calculated. As a result, it has been found that the loss peak of the blades of the rear loaded type is smaller than that of the front loaded type, as shown in FIG. 11 of the accompanying drawings. This is because the suction surface of the blades of the front loaded type is long downstream of the throat of the rotor blade outlet, so that the boundary layer is developed greater than with the blades of the rear loaded type. It is known that in the middle of the blades along their height, the blades exhibit intermediate characteristics between those at their base and tip.
From the above results, it can be understood that turbine blades which can suppress a secondary flow and suffer a smallest energy loss are of the front or intermediate loaded type at their base and of the rear loaded type at their tip. The inventors have designed a turbine having such characteristics.
The above and other objects, features, and advantages of the present invention will become apparent from the following description when taken in conjunction with the accompanying drawings which illustrate a preferred embodiment of the present invention by way of example.